Refrigerant flow control employing plural valves

ABSTRACT

A mechanical refrigeration system having flow metering devices therein for causing a balance between the boiling capability of the evaporator and the pumping capability of the compressor, whereby the system is adapted for air conditioner usage subject to varying evaporator heat loads and varying condenser coolant action.

O Umted States Patent 1151 3,640,086 Brody Feb. 8, 1972 [54] REFRIGERANT FLOW CONTROL 1,719,073 7/1929 Mofily..... EMPLOYING PLURAL VALVES 2,071,935 2/ 1937 Moffly..... 2,564,421 8/1951 Carter [72] ohm 3,119,559 1/1964 Heldorn ..62/217 [73] Assignee: American Standard Inc., New York, NY.

Primary Examiner-Meyer Perlin [22] Flled 1970 Attorney-John E. McRae, Tennes I. Erstad and Robert G. [21] Appl.No.: 15,160 Crooks I 521 u.s.c1. ..62/2l0, 62/217, 62/225 [571 ABSTRACT 1 [51] Int. Cl ..F25b 41/00 A mechanical refrigeration System having flow metering [58] Field of Search ..62/217,224, 225, 204, 205, devices therein for causing a balance between the boiling 62/206 2l 1 236/92 capability of the evaporator and the pumping capability of the compressor, whereby the system is adapted for air conditioner [56] References cued usage subject to varying evaporator heat loads and varying UNITED STATES PATENTS condenser coolant action.

1,990,663 2/1935 Moffly ..62/204 9Claims,4Drawing Figures 2 CONDINSER PATENTEUFEB 8|972 7 3,640,086 sum 1 OF 2 CONDENSER Fllll.

FIE E INVENTOR.

Hseamr/V B9004 PISTON MOTIQN 45 5o 55 TEMP. F'

PATENTEDFEB smz 3,640,086

sum 2 or z $.L. s.v.

PRESSURE 25.1. F

F 5| E F I t c 5 l --C.T. I l

CONSTANT SUPERH EAT :2 GO fir-1 EH3 =4 INVENTOR.

HERBERf M 5170M REFRIGERANT FLOW CONTROL EMPLOYING PLURAL VALVES THE DRAWINGS FIG. 1 schematically illustrates a refrigeration system utilizing the invention.

FIG. 2 is a chart illustrating certain temperature-motion relationships of thermostatic power elements employable in the FIG. 1 system.

FIG. 3 is a chart illustrating pressure-enthalpy relationships involved in mechanical refrigeration systems.

FIG. 4 is a diagrammatic representation of a thermostatic expansion valve commonly used to control refrigerant flow in conventional systems.

FIG. I

FIG. 1 schematically illustrates a refrigeration system of the invention which includes a conventional motor-driven refrigerant compressor 10, hot gas compressor discharge line 12, refrigerant condenser 14, liquid refrigerant line 16, first thennostatically operated restriction means 18, refrigerant evaporator 20, suction line 22, and second thermostatically operated restriction means 24 in the suction line.

In the illustrated system the restriction means 18 is intended to restrict flow of liquid refrigerant through line 16 to maintain the evaporatorpressure and temperature substantially constant in spite of substantial variations in heat load on the evaporator. Restriction means 24 is intended to modulate flow of gaseous refrigerant to the compressor to maintain the suction line temperature substantially constant or within acceptable limits. The general aim is to provide a system of modulating valves that produces a balance between the boiling capability of the evaporator and the pumping capability of the compressor. The system is designed particularly for air conditioner applications which must operate under conditions of varying condenser temperatures, and varying evaporator heat loads.

RESTRICT OR 18 Restrictor 18 is illustratively shown to include a housing 26 having a fixed partition 28 which is equipped with a plurality of ports 30 for unrestricted flow of refrigerant from space 32 to space 34. Disposed on partition 28 is a thermostatic operator 29 which may be constructed similarly to the power device shown in U.S. Pat. No. 2,368,l8l. This operator includes a cuplike metallic casing 36 having an inturned flange 37 at its mouth for securement of the cup to a cylindrical piston guide 38. Casing 36 contains a mass of solid thermal expansion material 40 which may for example be a combination of low melting point waxes changeable to the liquid state in a range of temperatures between 40 and 50 F. The drawing illustrates the device at a temperature of 40 F., i.e., with material 40 in the solid state. Assuming the fluid in chamber 34 is at a temperature slightly above 40 F., material 40 will thermally expand to deflect the elastomeric diaphragm 41 and the elastomeric plug 42 leftwardly, thereby moving piston 44 leftwardly against the action of compression spring 46. Spring 46 is trained between a fixed partition 48 and the flange of a cuplike spring retainer 50 carried by the piston 44. Retainer 50 is free to slide along the outer surface of member 38.

As piston 44 moves leftwardly its modulating valve element 43 moves toward a flow opening 45 in disc 47, thereby reducing the size of the flow path provided by the opening. In its FIG. 1 position disc 47 is held against a fixed partition 48 by means of a compression spring 49. During operation of the system valve element 43 does not fully engage disc 47; the disc therefore is held on the partition. As element 43 moves toward opening 45 refrigerant flow through opening 45 is progressively reduced directly in accordance with the temperature increase in material 40. Temperature decrease in material 40 produces a thermal contraction of the material, thereby allowing spring 46 to return piston 44 and valve element 43 rightwardly to their illustrated positions. thus permitting a progressively increased flow of refrigerant through flow opening 45. It is contemplated that valve element 43 will modulate back and forth to maintain the chamber 34 temperature within about l of some specific value, as for example 43 F. It is contemplated that the invention can be practiced using different temperatures and temperature ranges, depending on the refrigerant and environment.

The expansion material 40, being a solid material (as opposed to a gaseous material) is capable of developing very high forces on diaphragm 40 and piston 44 during the transition to the liquid state, for example pressures in excess of 1,000 p.s.i. Spring 46 merely acts to return piston 44 rightwardly against certain minor forces, principally the frictional resistance between sealing plug 42 and piston guide 38. On the expansion stroke of material 40 the potential internal forces are so great that the thermostatic operator is insensitive to the pressure in outlet space 34 and inlet space 31. Thus, the pressure drop across port 45 has no effect on the thermostatic operator or the position of the controlling piston 44; the position of the piston is controlled solely by the temperature of material 40. The thermostatic operator is therefore suited to control the position of the piston without regard to variations in pressure at different points in the system, as for example in liquid line 16 and/or suction line 22.

FIG. 2

FIG. 2 illustrates generally the type of temperature-motion curves obtainable with power elements of the type shown at 29 in FIG. 1. Assuming the power elements has a temperature of 40 F. when the parts are in the FIG. 1 position, the piston 44 would be at point A on the FIG. 2 chart; the piston valve element 43 would exert substantially no restriction of the flow through port 45. Temperature increase of material 40 to 45 F. would cause valve element 43 to take a position B (FIG. 2) substantially closing port 45. In the temperature range between 40 and 45 the valve element would move toward and away from the port opening 45 to thus exert a variable throttling action on the fluid.

During off periods the temperature of chamber 34 tends to equalize with ambient 33. Material 40 thereby expands to move the piston 44 leftwardly, thereby causing valve element 43 to force disc 47 leftwardly away from partition 48 against the action of spring 49. When the temperature of material 40 reaches about 52 F. the valve element 43 assumes a position C (FIG. 2); in this position disc 47 is spaced away from partition 48 a sufficient distance to define an extensive flow path around the disc periphery. This flow path is'sufficient to permit equalization of pressures across the system, as during cornpressor off periods. As will be seen from FIG. 2, temperature increase in material 40 above 52 F. produces substantially no added movement of the piston. This is not particularly significant in the case of restrictor 18, since added movement of the piston and disc 47 is only overtravel to preclude stress between the power element and its support wall 28.

RESTRICTOR 24 (FIG. 1)

Restrictor 24 may be constructed similarly to restrictor 18. Therefore similar reference numerals are employed where applicable. Restrictor 24 is shown with its piston 44' extended from power element 29'. During normal run periods the expansion material within power element casing 36' will be at a higher temperature than material 40 in housing 36, in accordance with the superheat produced in coil 20. This superheat may be on the order of 5 to 10 F.

The superheat causes piston 44' to normally hold its disc 47' away from partition 48', thereby providing a substantially unrestricted flow path through flow opening 51'. Should the superheat decrease the expansion material within housing 36' will contract to allow spring 46 to return piston 44' rightwardly, thereby permitting disc 47 to partially or wholly throttle the flow through opening 51'. This action tends to maintain the superheat within close limits. Additionally the thermostatic throttling action serves to meter gas flow to the compressor to reduce thecompressor inlet pressure, thus reducing the pumping capability of the compressor in accordance with reduced boiling capability of the evaporator. This action tends to reduce operating costs.

FIG. 1 illustrates the system under low evaporator load conditions; device 18 may then conceivably drop to a temperature of about 41 or.42 F., and device 24 may then be at a temperature of about 48 F. Under these conditions disc 47' exerts a throttling action on gas flowing through opening 51', thus raising the evaporator pressure and temperature, and also decreasing the boiling capability of the, evaporator. Additionally disc 47' restricts flow to the compressor, thus reducing the compressor inlet pressure and compressor pumping capability. The evaporator boiling capability decrease is thereby matched with a corresponding compressor pumping capability decrease.

Increased evaporator temperatures cause thermal expansion of material 40 in power element 29, thereby causing piston valve 43 to move leftwardly from its FIG. 1 position, thus throttling. the flow through orifice 45. The reduced refrigerant flow brings the evaporator pressure and temperature back down for attainment of higher boiling capability. It also tends to raise the superheat, which in turn causes power element 29' to move disc 47' away from opening 51'. More gasis then allowed to flow to the compressor so that higher compressor inlet pressure are produced. The result is increased compressor pumping capability to match the increased evaporator boiling capability. This modulating action of the two restrictors l8 and 24 continues throughout the normal run periods. The thermal substance in operators 29 and 29' may be the same material, the essential difference being that operator 29 operates in a low temperature range, for example between A and B in FIG. 2, while operator 29 operates in a higher temperature range, for example between B and C in FIG. 2.

FIG. 3

FIG. 3 is a pressure-enthalpy chart depicting the general operation of a refrigeration machine or system. Line S.L. represents the saturated liquid line, and line S.V. represents the saturated vapor line. Fluid to the left of line S.L. is liquid, fluid to the right of line S.V. is vapor, and fluid between lines S.L. and S.V. is vapor-liquid mixture. In an illustrative system condensed liquid refrigerant at A expands adiabatically across an expansion valve to a lower pressure B. vaporization in the evaporator takes'place at constant pressure along line B-C. Compression of the vapor in the compressor takes place along line C-D, and condensation of the vapor in the usual aircooled or water-cooled condenser takes place along line DA. In such a cycle the useful cooling is measured by line B-C.

The useful cooling may be somewhat increased by allowing the evaporator to produce some superheat. Thus, if the evaporator produces a few degrees superheat to condition C the useful cooling may be presented by line B-C'. The complete cycle would then be ABC'-D'-E-A.

The useful cooling may also be somewhat increased by allowing the condenser to subcool the refrigerant, as to condition A. The useful cooling would then be represented by line B-C, and the complete cycle would be represented by A'B -C-D --A'.

The pressure level in the evaporator is commonly controlled by a thermostatic expansion valve of the type schematically shown in FIG. 4. As there shown, the valve includes a housing 60 having an inlet 62 receiving liquid refrigerant from the condenser and an outlet 64 discharging expanded refrigerant to the evaporator. The amount of refrigerant is regulated by a valve element which is carriedby a stem 68 depending from a diaphragm 70. The space 71 above the diaphragm communicates with a capillary 72 leading to the charged bulb 74 which is clamped to the suction line of the system. Increase in superheat causes the bulb charge to produce an increased pressure in space 71, thereby causing the diaphragm to move the valve element downwardly against the action of spring 76. The downward force on the diaphragm is balanced by upward forces which include the pressure in space 78, spring 76, and the refrigerant supply pressure in chamber 63. The various pressures in practice vary somewhat depending on conditions within the condenser and evaporator. Hence, while the control by bulb 74 may be acceptable as concerns maintenance of superheat, yet it may not be acceptable as regards maintenance of a satisfactory evaporator pressure.

In FIG. 3 dotted lines C.T. represent lines of. constant temperature. Control of superheat by conventional bulb 74 can produce different evaporator pressures intersecting the respective line C.T. These different evaporator pressures produce lesser cooling than that denoted by line B-C' because the latent heat of vaporization at the lower pressures is less than at the illustrated pressure, and the lost cooling effect during flow across the expansion valve (as denoted by line F -B) is greater. In general it is preferable to carry out the refrigerant vaporization at as high a pressure as possible consistent with other factors such as the capacity of the compressor and the requirement for pressure head to move fluid from the high side to the low side of the system.

It is believed that the combination of metering devices 18 and 24 shown in FIG. 1 is preferable to the conventional bulb control in that it will more effectively maintain ,a relatively high and uniform evaporator pressure; this should provide greater cooling for a given size of equipment. Also, it is believed that the operating cost may be lessened because of the lessened refrigerant mass to be circulated and the lessened pressure difference to be maintained by the compressor.

As previously noted, device 18 meters flow of refrigerant to the evaporator so as to maintain a constant evaporator temperature and pressure. Device 18 may perhaps be visualized as a device for locating or stabilizing point B in FIG. 3. Device 24 may perhaps be visualized as a device for locating or stabilizing point C in FIG. 3. The intent, is that the system will produce cooling along the same pressure line B-C, irrespective of the evaporator heat load. Variations in cooling load will result in variations in amount of refrigerant pumped, but hopefully the metering action will maintain a balance between the evaporator boiling capability and compressor pumping capability as will result in a constant system cooling cycle'line B-C With such an arrangement it is believed possible to achieve modulated air conditioning solely by control of evaporator fan motor speed, fan blade pitch control, or other evaporator airflow controls.

The illustrated system has some additional advantages, including the ability to preclude compressor flooding. Thus, any flooding tendency causes power element 29' to move disc 47' rightwardly for throttling the flow through opening 51. Any liquid emanating from the evaporator encounters a pressure drop as it flows through opening 51', thereby reducing the suction line pressure for causing the refrigerant to flash vaporize. This action prevents compressor flooding.

The system is also advantageous in reducing evaporator frosting or freezeup, particularly under wintertime operation when air-cooled condensers may so effectively cool the compressed gas as to reduce the high side pressure below a value sufficient to adequately feed the evaporator. In conventional systems the evaporator tends to be so starved that the evaporator pressure and temperature drop enough to permit formation of frost on the evaporator fin surfaces. With the FIG. 1 control any excessive subcooling by the condenser will have no effect on the evaporator pressure because devices 18 and 24 are pressure-insensitive; i.e., they are influenced only by temperature. Devices 18 and 24 should maintain a substantially constant evaporator pressure and temperature irrespective of variations in condenser pressure.

The illustrated system may also be advantageous inachieving pressure equalization at each compressor shutdown. Thus at shutdown the temperatures of power elements 29 and 29' tend to equalize with the ambient 33. This ambient is in practice sufficiently above evaporator temperature that each power element will move its disc 47 or 47' leftwardly for rapid pressure equalization. This would in most systems be rapid enough to permit elimination of the conventional time delay devices which are commonly employed to keep the compressors off for predetermined periods, usually 5 minutes, after the compressor goes off.

Iclaim:

1. In a refrigerant system comprising a refrigerant compressor, refrigerant condenser, and refrigerant evaporator arranged in a closed loop circuit: the improvement comprising a first thermostatically operated metering means located between the condenser and evaporator, said metering means including a first thermostatic operator responding to evaporator temperature to maintain a substantially constant temperature therein; and a second thermostatically operated metering means located in the suction line between the evaporator and compressor, said second metering means including a second thermostatic operator responding to suction line temperature to maintain a substantially constant superheat; the thermostatic operator for each metering means being responsive solely to refrigerant temperature and being unaffected by refrigerant pressure.

2. The system of claim 1 wherein the first thermostatic operator is thermally engaged with refrigerant entering the evaporator.

3. The system of claim 1 wherein the first thermostatic operator comprises a contained mass of solid thermal expansion material changeable to the liquid state over a temperature range to maintain the evaporator at a first relatively low temperature; the second thermostatic operator comprising a contained mass of solid thermal expansion material changeable to the liquid state to maintain the second operator at a second relatively high temperature difierent than the first temperature.

4. The system of claim 1 wherein the first metering means comprises a thennostatic operator thermally engaged with refrigerant entering the evaporator, a port directing refrigerant across the thermostatic operator, and a valve element connected with the operator to regulate flow through the port; said operator and valve element being oriented to the port so that temperature decrease in the operator causes the valve element to enlarge the flow path through the port, and temperature increase in the operator causes the valve element to restrict the flow path through the port.

5. The system of claim 1 wherein the first metering means comprises a valve casing having a port-forming means therein defining a high pressure chamber communicating with the condenser and a low pressure chamber communicating with the evaporator; a valve controlling flow through the port; and

a thermostatic operator for moving the valve to thus regulate refrigerant flow through the port in accordance with the operator temperature; said thermostatic operator comprising a piston-container assembly disposed in the low pressure chamber so that refrigerant flows past the container as it moves toward the evaporator, and a mass of solid thermal expansion material contained within the container for producing an operating force on the piston; said expansion material undergoing a change to the liquid state over a range of temperatures spanning the boiling temperature of the refrigerant while in the evaporator.

6. The system of claim 5 wherein the port-forming means comprises a fixed partition having a relatively large pressureequalizer opening therethrough, and a floating member having a relatively small flow control port therethrough; said floating member being arranged to close said opening during normal run periods, and said thermostatic operator being effective to move the floating member away from the opening at shutdown of the system so that system pressures are equalized through the opening. Y

7. The system of claim 1 wherein the first metering means comprises a thermostatic operator thermally engaged with liquid refrigerant entering the evaporator; said operator having a force capability substantially greater than pressures existing in the high side of the system, whereby the operator is insensitive to system pressure variations.

8. The system of claim 1 wherein the first metering means defines a relatively small volumetric flow orifice when its ,operator is at evaporator design temperature, and the second metering means defines a relatively large volumetric flow opening when its operator is at design superheat.

9. The system of claim 1 wherein the second metering means comprises a valve casing having a port-forming means therein defining an inlet chamber communicating with the evaporator and an outlet chamber communicating with the compressor; a valve element controlling flow through the port; and a thermostatic operator for moving the valve to thus regulate flow through the port in accordance with the operator temperature; said thermostatic operator comprising a pistoncontainer assembly disposed in one of the chambers so that refrigerant flows past the container as it moves through the casing, and a mass of solid-liquid thermal expansion material contained within the container for producing an operating force on the piston due to change of state between the solid and liquid conditions; said operator and valve element being oriented to the port so that temperature decrease in the expansion material causes the valve element to restrict the flow path, and temperature increase in the expansion material tends to enlarge the flow path. 

1. In a refrigerant system comprising a refrigerant compressor, refrigerant condenser, and refrigerant evaporator arranged in a closed loop circuit: the improvement comprising a first thermostatically operated metering means located between the condenser and evaporator, said metering means including a first thermostatic operator responding to evaporator temperature to maintain a substantially constant temperature therein; and a second thermostatically operated metering means located in the suction line between the evaporator and compressor, said second metering means including a second thermostatic operator responding to suction line temperature to maintain a substantially constant superheat; the thermostatic operator for each metering means being responsive solely to refrigerant temperature and being unaffected by refrigerant pressure.
 2. The system of claim 1 wherein the first thermostatic operator is thermally engaged with refrigerant entering the evaporator.
 3. The system of claim 1 wherein the first thermostatic operator comprises a contained mass of solid thermal expansion material changeable to the liquid state over a temperature range to maintain the evaporator at a first relatively low temperature; the second thermostatic operator comprising a contained mass of solid thermal expansion material changeable to the liquid state to maintain the second operator at a second relatively high temperature different than the first temperature.
 4. The system of claim 1 wherein the first metering means comprises a thermostatic operator thermally engaged with refrigerant entering the evaporator, a port directing refrigerant across the thermostatic operator, and a valve element connected with the operator to regulate flow through the port; said operator and valve element being oriented to the port so that temperature decrease in the operator causes the valve element to enlarge the flow path through the port, and temperature increase in the operator causes the valve element to restrict the flow path through the port.
 5. The system of claim 1 wherein the first metering means comprises a valve casing having a port-forming means therein defining a high pressure chamber communicating with the condenser and a low pressure chamber communicating with the evaporator; a valve controlling flow through the port; and a thermostatic operator for moving the valve to thus regulate refrigerant flow through the port in accordance with the operator temperature; said thermostatic operator comprising a piston-container assembly disposed in the low pressure chamber so that refrigerant flows past the container as it moves toward the evaporator, and a mass of solid thermal expansion material contained within the container for producing an operating force on the piston; said expansion material undergoing a change to the liquid state over a range of temperatures spanning the boiling temperature of the refrigerant while in the evaporator.
 6. The system of claim 5 wherein the port-forming means comprises a fixed partition having a relatively large pressure-equalizer opening therethrough, and a floating member having a relatively small flow control port therethrough; said floating member being arranged to close said opening during normal run periods, and said thermostatic operator being effective to move the floating member away from the opening at shutdown of the system so that system pressures are equalized through the opening.
 7. The system of claim 1 wherein the first metering means comprises a thermostatic operator thermally engaged with liquid refrigerant entering the evaporator; said operator having a force capability substantially greater than pressures existing in the high side of the system, whereby the operator is insensitive to system pressure variations.
 8. The system of claim 1 wherein the first metering means defines a relatively small volumetric flow orifice when its operator is at evaporator design temperature, and the second metering means defines a relatively large volumetric flow opening when its operator is at design superheat.
 9. The system of claim 1 wherein the second metering means comprises a valve casing having a port-forming means therein defining an inlet chamber communicating with the evaporator and an outlet chamber communicating with the compressor; a valve element controlling flow through the port; and a thermostatic operator for moving the valve to thus regulate flow through the port in accordance with the operator temperature; said thermostatic operator comprising a piston-container assembly disposed in one of the chambers so that refrigerant flows past the container as it moves through the casing, and a mass of solid-liquid thermal expansion material contained within the container for producing an operating force on the piston due to change of state between the solid and liquid conditions; said operator and valve element being oriented to the port so that temperature decrease in the expansion material causes the valve element to restrict the flow path, and temperature increase in the expansion material tends to enlarge the flow path. 